Resiliently frictionally sway stabilized railway car

ABSTRACT

Swaying or rolling motion of a railroad car body is damped by a form of energy-absorbing coupling between each side of the trackfollowing portion of the car and the corresponding side of the spring-supported portion, providing substantial energydissipative force opposing approach motion at the respective sides of the car but with separation substantially unimpeded. Stabilization of the body against oscillatory motion at the resonant frequency of the springs is obtained without substantially increasing the hazard of derailment on curves in any range of speeds. A simple form of construction of an energyabsorber for installation on existing conventional cars is described.

0 United States Patent 1 1 1111 3,731,638 Tack 1 May 8, 1973 54 RESILIENTLY FRICTIONALLY SWAY 3,482,530 12/1969 Tack ..105/199 R STABILIZED RAILWAY CAR 3,351,336 11 1967 Blake ...105/199 R x 2,685,845 8/1954 Gassner ...lO5/l99 R X Inventor: Carl Tack, 157 Under! 2,961,974 11 1960 Seelig .105 453 x hurst, 111. 60126 [22] Filed: Jam 28 1970 Primary Examiner-Robert G. Sheridan Assistant ExaminerI-loward Beltran 1 1 pp 6,473 AttorneyLeonard G. Nierman 52 us. c1 ..105/199 A, 105/199 CB, 105/453, [57] ABSTRACT 303/138 Swaying or rolling motion of a railroad car body is [51] Int. Cl. ..B61f 5/14, B61f 5/24, F160 17/04 d m ed by a form of energy-absorbing coupling [58] Field Of Search ..105/199, 453, 193, between each ide of the traclpfollowing portion of 105/197 197 199 199 199 199 the car and the corresponding side of the spring-sup- 199 138 ported portion, providing substantial energy-dissipative force opposing approach motion at the respective [56] References Cited sides of the car but with separation substantially unimpeded. Stabilization of the body against oscillatory UNITED STATES PATENTS motion at the resonant frequency of the springs is ob- 1,347,898 7 1920 Eaton ..105/199 R mined Without Substantially increasing the hazard of 2,352,039 6/1944 Travilla, Jr.... ..105/199 R derailment on curves in any range of speeds. A simple 3,020,857 2/1962 Dean ..105/193 X form of construction of an energy-absorber for instal- 3,043,241 62 Ortner 1-- 199 CB lation on existing conventional cars is described. 3,415,203 12/1968 Hughes et a1. ..105/199 R 3,459,139 8/1969 Love ..308/138 X 13 Claims, 11 Drawing Figures PATENTED 81m 3.731.638

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SHEET 0F 4 ax-mbia INVENTOR. CARL E. TACK ATTORNEY RESELIENTLY FRICTIONALLY SWAY STABILIZED RAILWAY CAR The subject matter of this application is in part disclosed in an application of the same inventor filed Aug. 26, 1969, Ser. No. 853,176, now abandoned, and a continuation thereof filed Jan. 20, 1971, Ser. No. 210,175.

This invention relates to the stabilization of railroad cars, and particularly freight cars of relatively long length, against lateral swaying or rolling motion at the resonant frequency of their spring suspension, and more particularly to structure for energy-absorbing damping of such lateral oscillatory motion which avoids the introduction of deficiencies of performance introduced by prior art devices for the same purpose.

The basic construction of a freight car essentially universally employs a car body supported at each end by a truck having wheels (normally four on two axles) mounted in laterally opposed side-frames, in which are mounted compression springs supporting the bolster portion of the truck, upon which the bolster of the car body is pivotally supported. The springs are primarily employed to isolate the body from vertical motions of the wheels of a sharp impact nature, such as those occurring at crossings and the like, and of a rapid oscillatory nature such as on rough track at high speeds. As in the case of any other spring suspension, such a system inherently has a resonant frequency. If rolling or rocking force is periodically applied to one or both sides of such a mechanical system with a periodicity in the region of the resonant frequency of the laterally opposed springs, there results a swaying or rolling action of large oscillatory amplitude. As hereinafter pointed out and explained in greater detail for full understanding of the invention, this phenomenon has introduced serious problems in the use of freight cars of relatively great length as compared with the shorter lengths which were at one time standard. For the reasons later explained, operation of cars of particular lengths equipped only with conventional trucks at certain speeds can produce wild side-to-side rolling or swaying of a prohibitive amplitude.

To deal with this problem, there have heretofore been proposed, and in some cases there are now in use, various devices to provide damping of such side-to-side rocking or swaying of a car body. Prior to the present invention, one of the most satisfactory of these, so far as is known, was the construction of the present inventor disclosed in U. S. Pat. No. 3,482,530. Upon extensive testing, however, it was found that this construction, although more satisfactorily solving the problem for high-speed operation of long cars than earlier devices, introduced an undesirable incidence of derailment on banked curves at relatively slow speeds on track in marginal condition of maintenance.

The present invention flows from theoretical and experimental observation of the dynamic forces and motions involved in the operation of a car, and lies in the provision of a construction which satisfactorily solves the problem of control of swaying at all speeds, without substantial impairment of the ability of the spring system to prevent transfer of impact-type wheel excursions to the body, and without derailment hazard.

Briefly stated, the construction of the present invention in its broader aspects, like that of the patent of the same inventor just mentioned, employs a direct coupling of each side of the spring-supported structure to the corresponding side-frame of the supporting truck. In the present invention, this coupling offers relatively high, but frictionally yielding, resistance to downward motion of the side of the car body toward the corresponding side-frame but little or no resistance to upward or separative movement from the normal or average relative position.

It is a further principal object of the invention to provide such a construction in which the coupling is directly applied to the portion of the spring-supported structure which is horizontally pivoted with respect to the side-frames, while at the same time avoiding any undue interference with freedom of relative rotation between the trucks and the body.

Another frequently-observed undesirable phenomenon in car operation is the spasmodic rapid horizontal oscillatory chattering of the wheels often called shimmy, which occurs because the wheel flanges are sufficiently inward of the rails to permit a small angular freedom of the track-following assembly. Under certain conditions, there is established a vibratory motion wherein the flanges of first one set of diagonally disposed wheels and then the other rapidly alternate in striking the rails and bouncing to the opposite position. This continues, once established, until sufficient change of speed, rail spacing, etc., occurs to eliminate it. The present invention provides a frictional damping impedance to such motion which substantially reduces the incidence and duration of wheel-shimmy.

In addition to accomplishment of the principal objects above described, the present invention in its narrower aspects further implements the general type of construction just summarized with simple and convenient construction features for a stabilizer or strut" providing the requisite action and readily adapted for installation on conventional cars.

More detailed explanation of the invention, both in its broad and narrow aspects, will be best understood with the aid of schematic illustrations provided in FIGS. 1 to 3 of the annexed drawing for this purpose. Accordingly, further discussion of the aims and objects of the invention, and the manner in which they are accomplished, is here deferred for consideration along with the drawing, in which:

FIG. 1 is a highly schematic illustration of the basic elements of a conventional freight car as related to the present invention;

FIG. 2 is a schematic illustration of a railroad track with alternated or offset track-joints of a type in common use;

FIG. 3 is a schematic sketch similar to FIG. 1, but illustrating the conditions which prevail when a car proceeds along a banked curve at a speed far below the speed for which the banking is designed;

FIG. 4 is a view in front elevation, partially in section, showing a railroad car incorporating the present invention;

FIG. 5 is an isometric perspective view of a portion of the construction of FIG. 4 consisting of the sideframes of the truck, the bolster portion of the body and energy-absorbing devices mechanically coupling these elements;

FIG. 6 is a more or less schematic plan sectional view taken along the line 6-6 of FIG. 4;

FIG. 7 is a fragmentary side elevation further illustrating the manner of coupling of the side-frames to the car body and the energy-absorbing structure;

FIG. 8 is a view in longitudinal section of the internal construction of the energy-absorbing device;

FIG. 9 is a transverse sectional view taken along the line 9-9 of FIG. 8; and

FIGS. 10 and 11 are fragmentary views corresponding to FIG. 7 but showing other constructions embodying the more basic teachings of the invention.

In the schematic illustration of FIG. 1, as will be readily seen by those skilled in the art, and as will in any event be later apparent, the actual location of the parts of 'a conventional railroad car has been substantially distorted for simplicity of understanding of the problems which the present invention, unlike certain devices of the prior art which have been found lacking, is found to solve satisfactorily. As shown in FIG. ll, the car body 10 is vertically supported on the truck bolster 1 1 by a body bolster 12 with a pivotally engaged center plate 13. There is a substantial gap or clearance at the side-bearings 14 in most car constructions. The depending round protrusion forming the center plate 13 fits rather loosely into the corresponding hollow in the bolster 11 and thus permits relative rocking motion between the body 10 and the truck bolster lll under conditions of large tilt although the entire weight of the body is otherwise wholly supported by the center plate 13.

The truck bolster ll is supported by the springs 18 and 20, which are in turn supported by the track-fob lowing assembly 22 (schematically shown as merely the axle) upon which are mounted the wheels 24 and 26.

The springs 18 and 29 are of course substantially matched, and the mounting of the body it) accordingly has one or more resonance frequencies of vibration in at least two modes, vertical and sidewise or rocking, differing in relative phase of oscillations of the springs on respective sides. It will be seen that if the wheels 24 and 26 simultaneously encounter a drop or an elevation of their respective tracks, vertical vibration is stimulated at the natural frequency of the system. it" one wheel encounters an elevation while the other encounters a drop, the induced vibration is pendulum-like. if only one wheel encounters an elevation or drop, the vibrations of the springs 18 and are neither wholly in phase nor wholly out of phase, and the center of gravity of the loaded car is stimulated to vibrate in a combination of vertical and swaying (pendulum-like) motion.

Car trucks are of course always designed with such considerations in view, and springs such as those shown schematically at 18 and 20 are designed to have a sufficiently low resonant frequency to isolate the car body from the impact-type shock at crossings, etc. for which the spring support is primarily designed, vertical vibration so induced being harmless.

in FIG. 2 is shown in schematic fashion a typical length of railroad track 28, with track joints shown at 30. The joints 30 are staggered, those on each track being approximately opposite the section midpoints of the other track. in non-welded track, each track section is somewhat firmer at its median portion than at the joints, each wheel of a car normally following a more or less sinusoidal path of amplitude varying with track maintenance condition. As the car of HG. ll

proceeds along a straight track, the track-following assembly 22 accordingly undergoes a slight rocking action with each of the wheels 24 and 26 being alternately higher and lower than the other. in many modern freight cars, the distance between trucks is of the same order as the length of rails in common use. When a car of such length is operated on standard staggered track such as shown in lFlG. 2, the rocking or rolling action of the front and back trucks occurs in complete synchronism. As such a car is run over a track at any constant speed, there is thus produced a periodic synchronous rocking or rolling of the trucks. At very low speeds, this results merely in a wholly corresponding slight rocking motion of the car body. However, as the speed is increased there is reached a condition where the speed of rocking of the trucks corresponds with the resonant frequency of the oscillating system of FIG. i, this being called the critical speed. At this speed the resonant spring system is excited or driven at its natural frequency of vibration, and the swaying or rolling excursions of the car body become much greater than the alternative excursions of the trucks. The running of such a car over any substantial length of standard ofiset or staggered track can readily produce extreme swaying or rolling of the car body. Particularly in cars (or loads) of great height, such as modern automobile transports, such operation is wholly unacceptable.

The resonant frequency of rolling or rocking action of a car is a function both of the design of the car and its load. Typical resonant rocking frequencies of modern long freight cars are of the general order of one-half cycle per second or 2 or 3 cycles per second. With standard 39 foot track length (and truck spacing) this corresponds to speeds from as low as 12 or 13 miles per hour to speeds of 50 or miles per hour. Since freight trains, as a matter of practical necessity, inevitably include a large variety of cars and loadings, virtually any freight train having numerous freight cars of length comparable to track-section length would have at least a few cars which would sway to and fro to a dangerous degree at virtually any running speed were no further provision made.

A variety of devices have been provided to deal with the problem of swaying. These range from conventional oscillation-damping devices such as hydraulic shock-absorbers to more refined constructions for this specific purpose such as that described in the patent of the present inventor earlier mentioned. Conventional motion-damping devices are often incorporated in the trucks, between the side-frame and the truck bolster. The more desirable constructions, however, employ a direct damping coupling between the body and the track-following portion of the truck, by passing the truck bolster as indicated schematically by the dotted lines 32 and 34 in FIG. 11. Such direct energy-dissipative coupling between the body lit) and the track-borne assembly 22 is particularly desirable in car constructions wherein there is a substantial gap or clearance 14 between the body bolster portion and the truck bolster portion because of the great addition to oscillatory swaying which is produced in operation at the resonant frequency when the car body is free to tilt with respect to the truck bolster. it will be seen that the mechanical system consisting of the body structure it) and the truck bolster 11 has three distinct conditions of stability. The flat under-surface of the center plate 13, resting in the corresponding recess in the truck bolster l1, maintains a constant relative positioning (assuming a reasonably balanced load) when the spring-supported portion is in a wholly vertical position or slightly tilted. However, if the tilt reaches a point where the center of gravity of the body is laterally outward of the center plate, the entire body immediately shifts to the position wherein the body is partially supported laterally outward of the center plate (see F IG. 3, later to be discussed in another connection). It is in principle, of course, possible to prevent the abrupt shifting by providing bearing structures which eliminate the gap or clearance shown at 14, but most practical car constructions, although using lateral bearings as shown in FIG. 1, nevertheless permit a substantial abrupt shift when a high degree of tilt is reached. The rapid additional acceleration of the load thus produced at a point of time when the underlying spring is already being compressed adds greatly to the amplitude of the excursions which occur, once they are of sufficient amplitude to produce the undesired stepwise rocking motion of the car body with respect to the truck bolster. in addition, of course, such freedom of relative motion between the body and the truck bolster produces highly undesirable jarring of the load. 7

Despite such shortcomings, damping devices in most common use are incorporated wholly in the truck to avoid the problems of providing a mechanical linkage permitting free pivoting of the truck with respect to the car body. Attempt is often made to provide sufiiciently high damping to prevent reaching oscillation of sufficient amplitude to unseat the center plate except on the roughest track (i.e., track with greatest yielding at the joints). Such damping obviously involves considerable sacrifice of isolation of the load from vibration, and the best possible compromise is made in each case. In addition to reducing load isolation, the introduction of high damping with types of damping devices heretofore employed introduces derailment problems to be later described.

In the patent earlier referred to, and certain other prior art, there is employed the general type of damping shown at 32 and 34 in FIG. 1, i.e., directly between the car body and the corresponding side frame. Although such approaches to the problem have been demonstrated to be capable of excellent sway stabilization without great sacrifice of the effectiveness of the springs in isolating the load from the impact-type jarring and vibration which is the basic function of the springs, there are found to be introduced occasional derailments on curves, for reasons now to be discussed, also applicable to many constructions wherein the stabilization is provided wholly in the truck.

In FIG. 3 is illustrated (with some exaggeration) the condition which prevails when the structure of FIG. 1 proceeds along a banked curve at a speed slower than the speed at which the banking angle compensates for the effects of centrifugal force. The rocking motion of the axle and frame assembly 22 is essentially the same as in the case of straight track, except that it of course occurs about a median position of tilt. Because of the shift in the center of gravity, the spring-supported portions have a tilted position with respect to the track-following assembly, the inner or lower spring 13 being somewhat compressed and the outer or upper spring 20 being somewhat extended. In the schematic illustration, the body 10 is shown in one of its two sideways tilt positions with respect to the truck bolster 1 1. The oscillation imparted by the rocking of the track-following assembly 22 thus occurs about this non-equalized position. If the amplitude of oscillation becomes sufficiently large, particularly with loads having a high center of gravity, not only may the lower spring 18 periodically bear the entire load, but indeed, particularly with cars having a high center of gravity, the upper or outer portion of the spring-supported structure may exert an upward force. If there is substantial friction in the coupling 34 (or in a damping device wholly on the truck), the upper or outer wheel may be lifted from the track to a point where its flange becomes disengaged from the track. Due to limitations of precision of track spacing and other factors, when the wheel then returns to the track, it may do so in a condition wherein the flange is seated on the upper track surface, whereupon derailment is a frequent prompt consequence.

The problem thus presented is actually more complex than described above. There are possible a number of resonant frequencies corresponding to various modes of vibration and harmonics which may be excited when a particular car is operated over a wide range of speeds, magnitude and balance of loads, and tracks of varying condition. Constructions which may appear to solve the problem at the nominal critical speed are empirically found to demonstrate problems of detailing and the like at much higher speeds due to difiering vibrational loads and harmonic excitations of the overall spring systems.

Previous approaches to the problem have either involved such cost and complexity as to be prohibitive, or have involved performance sacrifices which are avoided by the present invention. The reason for the impaired performance of previous constructions is found to lie in the force-transfer characteristics of the constructions heretofore employed. In most cases, such as hydraulic damping devices often used, the energyabsorbing frictional or viscous opposition to relative motion is a function largely of the relative velocity of the parts coupled. In another type of energy-absorbing damping, friction is essentially constant at all times. Damping devices of various other characteristics have also heretofore been employed. In the present invention, the damping force or frictional energy-absorption which has been found to produce a greatly superior performance may be briefly characterized as follows: The damping or energy-absorbing frictional force is applied at each side of the car only during motions wherein the car body is moving toward the track-following portion at the corresponding side, providing no resistance to motion in the opposite direction, i.e., in the direction of separation. With such a characteristic, oscillatory motions of both a vertical nature (bouncing) and of a lateral nature (rocking) are rapidly damped without production of any possibility of wheellifting. In the preferred construction of which embodiments are below described, this characteristic is obtained by mounting the energy-absorbing device on the body (or on the side-frame) and providing a movable coupling member which follows separative motions without opposing them, but which opposes motion in the opposite direction, frictionally yielding to absorb the energy of such approach motion. As a further advantageous feature of construction, the frictional force opposing motion in the latter direction increases with downward position of the side of the spring-supported portion, thus providing much greater damping between the side-frame and the corresponding side of the carbody in the region of very high compressions of the corresponding spring than in the normal or median position, so that large-amplitude oscillations are highly damped without introducing objectionable transfer of small-amplitude excursions of wheels to the car body.

Referring first to the embodiment shown in FIGS. 4 through 8, the correspondence to the schematic drawing of FIGS. 1 and 3 will readily be recognized. The wheels 36 and axle 38 non-resiliently support, by conventional journal boxes (not shown), conventional side-frames 40 in which are mounted the springs 42 atop which rest the ends of the truck bolster 44, extending through apertures 46 in the side frames. The upper surface of the truck bolster 44 has a conventional round recess or truck center plate 48 receiving the conventional body center plate 50 formed on the lower surface of the body bolster 52. A web portion 54 extends across the lower end of the body of the car and a shear panel 56 extends out from the lower edge of the vertical web portion. The center sill 58 extends longitudinally forward and rearward of the body bolster, with which it is integral. The truck bolster has a pair of bearing members 60 laterally outward from the center and the body bolster 52 has bearing members 62 correspondingly located to form side-bearings 64 aiding in support of the body in a tilted condition, but at all times permitting horizontal rotation of the truck relative to the body.

The elements thus far described, and their interconnection, are in all respects wholly conventional and are illustrated more or less schematically and not further described.

Attached, preferably by welding, to the web 54 above the upper surfaces of the side frames 40 are brackets 65 mounting energy-absorbers generally indicated at 66, the internal construction of which will be later described. The movable portion of each energy absorber 66 constitutes a rod 68 protruding region, these are joined, preferably by welding, to four internal plates which thus serve as a portion of the housing and afford internal friction surfaces or areas.

At the top, the angle members 82 are welded to an indownwardly through an aperture 70 through the shear panel 56 in alignment with the upper surface of the corresponding side frame 40 (when the truck is in the position of straight motion). At the lower end,each rod 68 is pivotally pinned at 72 to the end of a crank lever 74.

The crank lever 74 (as best seen in FIG. 7) has a lower surface curved in the vertical plane, having its lowest point or bearing surface 76 laterally opposite the center of the center-plate 50, i.e., at the front-to-rear midpoint of the sideframe 40. The opposite end of the crank 74 is bifurcated to form arms 78 each pivotally mounted by ears 80 affixed to the undersurface of the shear panel 56. The central portion of the upper surface of each side frame 4% has a bearing plate 81 thereon, with which the bearing portion 76 of the crank 74 makes sliding contact.

The internal construction of the energy-absorber 66 is shown in FIGS. 8 and 9. The housing or casing is formed by four corner angle members 82. in the lower ternally seated end-cap 86 having a central aperture 9t).

Loosely mounted around the rod or shaft 68 are oppositely disposed wedge rings 92 and W3. Each wedge ring has a central aperture 96 of tapered diameter, being largest at the facing surfaces and smallest at the opposite surfaces, but passing the rod freely and loosely at all points to accommodate the slightly arcuate motion of the lower end of rod 68 to be later described. The periphery of each wedge ring is substantially square and in the general shape of a truncated pyramid, each side of the square having the outer surface bevelled to form a wedge of minimum dimensions in the inner or facing region and of maximum dimensions in the opposite region. Between the wedge surfaces of the rings 92 and M are disposed friction shoes 98 each of L-shaped cross-section and having one arm in sliding engagement with one of the plates 84 and the other arm in sliding engagement with an adjacent plate 84. The upper and lower edges of the shoes 98 are bevelled in correspondence with the bevelling of the wedge rings, which are in contact therewith. A flange or collar 99 formed on, or affixed to, the outer portion of the rod 68 seats the outer surface of the ring 94, with a suitable washer 100 interposed. A coiled spring 102 surrounds the inner end of the rod 68 and is compressed between the surface of the ring 92 and a disk flange or washer 104, which is adjustably positioned by a nut 106 on the threaded inner end of the rod 68.

A coiled spring 1, of much larger magnitude and capacity than the spring 102, surrounds the entire upper portion of the movable assembly just described, being under compression between the end cap 86 and the laterally outer portion of the inner wedge ring 92, which is formed with a rib separating the radially inner seating of the end of spring 102 from the radially outer seating of the surrounding heavier spring 108.

The energy-absorption characteristics of the construction shown in FlGS. and 9, constituting a friction strut, may be understood from consideration of the construction. The moving portion of the assembly slides in the housing with a frictional force controlled by the lateral force spreading apart the shoes 98, which is in turn controlled by the magnitude of the spring forces urging the wedge rings 92 and 94 together. The minimum or biasing force is provided by the small spring 102. This spring will be seen to exert no longitudinal force urging the rod 68 in either direction, and the lateral force exerted on the shoes 98 by this spring is more or less uniform at all times.

The larger spring 1108 serves a dual purpose, both biasing the movable assembly outward and exercising variable control of the frictional resistance to motion. The friction established by the light biasing spring 102 is relatively small and is effective primarily when the rod 68 is in a wholly extended position with the spring 108 under substantially no compression, i.e., with no inward (upward in FIG. 8) pressure exerted at the outer-end pivot point '72. The position assumed by the rod in this condition of no inward force is determined by the adjustment of the nut 11, which is usually adjusted to sufficient tightness so that lifting of the lever from engagement with the side-frame which normally supports it does not produce removal of the shoes 98 fromthe housing. If so desired, a pin or similar stop-member (not shown) may be inserted across the housing at the outer end to limit the outward motion and thus permit adjustment of the nut 106 to be made without restriction to such a minimum friction requirement if it is desired to make the first portion of an inward stroke of the rod '68 essentially entirely frictionless. In such event, the adjustment of the nut 106 can be used as a trimming adjustment, to provide a preestablished frictional damping or opposing forceto inward motion of the rod 68 at any predetermined inward position without any restriction to minimum friction at the outermost position.

With the internal construction of the energy-absorbers 66, shown in FIG. 8 and 9, understood, the motions and forces produced thereby in the coupling between the body and the side frames earlier described may be readily comprehended. As best seen in FIG. 6, the lowermost portion 76 of the crank lever 74 is in frictional sliding contact on the bearing plate 81 atop the corresponding side frame 40 leaving the truck free for pivotal motion at curves without the use of universal joints or any similar constructions introducing problems of complexity and mechanical failure. The friction, at the same time, reduces the susceptibility of the side-frame structure to shimmying oscillation in the horizontal direction. The energy-absorbers 66 are installed at a height such that with the car stationary on a level track, the rods 68 are driven by the weight of the car body part-way into the housing. The degree of penetration in this median or normal position will of course vary somewhat with the weight of the load in the car. It is of course possible to match a particular detailed design or adjustment to the car each time it is loaded (or emptied) but the mode of operation is to a large extent self-compensating for degree of loading because of inertia change. In this normal or median position, the compression of each large spring 108 is relatively small as compared with the degree of compression which occurs on alternate sides upon the occurrence of rocking relative motion between the body and the side frames.

When rocking motion occurs on a straight track, damping action occurs alternately on one side and then the other. As the rod is driven into the housing, the pressure of the spring 108 on the wedge rings 92 and 94 increases, thus increasing the friction between the shoes 98 and the plates 84 to dissipate oscillation energy of the springs 42. However, this frictional dissipation is essentially unidirectional. When the motion reverses, i.e., as the car body and the side frame commence to separate, the spring 108 forces the rod 68 outward to maintain the contact at 76. Obviously, the energy-absorber 66 presents no frictional resistance to this separative motion. On straight track in well-maintained condition, both levers are at all times in contact with their corresponding side-frame plates 81, oscillating, for example, between the positions shown in solid and dotted forms in FIG. 7. However, in the event of relatively large separative excursions, the rod 68 may reach the outer limit of its travel before the separative motion ends, in which event, as shown by the dotted indication of the lowered position of the side frame 40 in FIG. 7, contact may be broken, until the motion is arrested by the action of the absorber on the opposite side, and then again ire-established to commence the compression of the spring 108. As later pointed out, such adeparture from contact may also occur momentarily in the event a downward wheel excursion, even though of small magnitude, is too fast to be followed by the arm.

The absence of resistance to separative motion wholly eliminates the problem of raising forces on the upper wheel in proceeding around a curve at a speed substantially below the speed for which the banking compensates. In addition, the maximum safe speed of operation around a given curve is substantially increased, as compared with conventional damping constructions.

When the car is operated on a curve at relatively low speed, the swaying or rocking oscillations occur about a static or bias condition in which the springs 108 on the lower and upper sides are respectively compressed and extended, thus producing greater frictional action on the lower side than on the upper side. The mean position in the oscillatory motion of the body with respect to the wheels thus produced is somewhat more erect than the static position assumedwhen the car is stopped. Similarly, at speeds sufficiently high to produce (were the track welded) constant outward tilt of the car body, the damping action occurs predominantly at the outer side, thus again making the median position during oscillation somewhat more erect than previously known energy-absorbing provisions. Irrespective of speed, the amplitude and maximum velocity of swaying motion is greatly reduced without introducing any hazard of derailment such as heretofore produced in constructions capable of comparable anti-sway performance.

The damping provision of the invention, in addition, produces a minimum of interference with the action of the springs in isolating the car body from impact-type rapid excursions of the wheels such as at crossings, sudden dips at very soft joints, etc., and in high-speed operation on rough track. Although the release of friction within the damping device is sufficiently fast so that swaying action (except of very large magnitude) does not produce disengagement of the arms or levers 74 from the bearing plates 81, small downward excursions and returns of the wheels, if sufficiently fast, are not followed due to the finite time required for release of friction within the damping device at the instant when the side frame suddenly drops down and also the minimum friction established by the bias spring 102. The greater the compression of the spring 108 at the time of such an occurrence, the greater the time lag before the lever 74 is able to follow such a rapid downward motion. Accordingly, the rapid upward return of the wheels produces minimal jarring of the load even when (as is uncommon) such a track discontinuity is encountered at an instant when the frictional force within the clamping device is very high. The minimum friction is desirably adjusted (or may less advantageously be fixed by suitable precise selection of the larger springs, if bias springs are omitted) so that moderately large body oscillations are wholly followed at the cars critical speed but separative motions occurring very much faster than this at either side quickly break the contact. The damping action is thus disabled during the return approach relative motion occurring immediately after such a separative relative motion.

Although of utility in cars of virtually any length, the structure is particularly advantageous in cars wherein the truck-to-truck spacing is greater than 33 feet, particularly in the range of 36 to 42 feet.

In FIGS. and I l are shown alternate constructions employing the energy-absorbing device of FIGS. 8 and 9 which produce essentially identical operation. In FIG. 10, the energy-absorber 112, of internal construction similar to that of those already described, is mounted on the side-frame structure 114 in inverted orientation as compared with the previous construction, and the crank or lever 116 makes sliding engagement with a bearing plate 118 mounted on the undersurface of the car body 120. In this case, the rod 122 which is the reciprocable portion of the energy-absorber 3112 is pivoted in the central region of the lever 1116, the end portion of the lever making the sliding contact in this embodiment. In FIG. 11, the construction is similar to that first described, the primary difference being that the crank or lever arm, here designated 124, is mounted in ears 126 outward, rather than inward, of the absorber 66 and makes contact with the bearing plate 81 at its free end in a manner generally similar to FIG. 10. It is important to note that in FIGS. 10 and II, as in the first embodiment, the sliding contact is made at apoint at which no lateral force or torque is exerted on the side-frames; in the first embodiment described and in FIG. 11, the sliding contact is directly at the laterally central portion of the upper surface of each side-frame, whereas in FIG. 10 this is accomplished by locating the contact point in vertical alignment with this portion of the side-frame. By such constructions there is avdided any lateral loading of the side-frames.

It will be recognized that the embodiments of FIGS. 10 and 11 are representative only of relatively minor variations of the embodiment more fully illustrated, and that the broader teachings of the invention may be practiced with a variety of internal constructions of energy-absorbers 66 or I12 providing the desired external characteristics, although the internal construction illustrated is particularly advantageous.

What is claimed is:

1. In a railroad car having a wheel-supported sideframe portion, a body portion, and a truck bolster supported by a spring on the side-frame portion and having the body portion pivotally supported thereon, the improved construction having energy-absorbing damping means at each side of the car, by-passing the bolster, for opposing approach motion of that side of the body portion of the car relative to the corresponding side of the side-frame portion of the car but substantially free of resistance to separative motion, the damping means having substantially greater energy-absorption of approach motion in the region of high compression of the spring than in the region of normal compression, and means to disable the damping means during return motion occurring immediately after separative motion substantially faster than critical-speed body-oscillation motion.

2. The car of claim 1 wherein each damping means includes means primarily responsive to instantaneous relative position of the side-frame portion and the body portion for producing a continuously increasing force characteristic during approach motion.

3. The car of claim 1 characterized by the damping means having a housing secured to one of said portions and a force-transfer member in sliding engagement with the other of said portions, said disabling means including means to adjust the maximum speed of following of separative motions by the force-transfer member.

4. The car of claim 11 wherein each damping means comprises a housing, an extending operating member having an inner portion movable in the housing, a spring within the housing, stressed by inward motion of the operating member, adjacent friction surfaces on the housing and the portion of the operating member within the housing, and means within the housing responsive to variation of stressing of the spring to vary the frictional engagement of the friction surfaces correspondingly.

5. The car of claim 4 wherein each damping means has an auxiliary adjustable bias spring coupled to the frictional engagement varying means.

6. The car of claim 5 wherein the bias spring is mounted on the operating member for motion therewith, so that the force exerted thereby is positionindependent.

7. In a railroad car having a track-follower portion and a body portion spring-supported on the track-follower portion for horizontal pivotal motion with respect thereto, the improved construction having energy-absorbing damping means at each side of the car opposing downward motion of that side of the body portion of the car relative to the corresponding side of the track-follower portion of the car but substantially free of resistance to upward motion, the resistance of the damping means to downward motion being substantially greater in the region of high compression of the spring than in the region of normal compression, the damping means having a housing secured to one of said portions and a force-transfer member in sliding engagement with the other of said portions, the forcetransfer member comprising a lever pivoted on the same portion to which the housing is secured, the damping means having an extending member reciprocable in the housing and pivotally secured to the lever.

8. The car of claim 7 wherein the housing is secured to the body portion and the lever has a portion slidingly engaging the side frame of the truck laterally opposite the pivotal support of the body portion.

9. The car of claim having the housing secured to the end of the body substantially above the bottom thereof, the lever having a portion pivotally mounted on the undersurface of the body and a depending portion slidingly engaging the side frame.

W. A motion-damping assembly for a railway car having pivotally connected truck and body portions comprising a housing and a fulcrum member both having means for attachment to the same portion of the car, an actuator extending into the housing, a lever pivotally connected to the actuator and the fulcrum member and having a surface constructed and arranged for sliding engagement with an adjacent surface of the other portion of the car, a spring urging the lever into such engagement, and direction-selective energy-absorption means in the housing including means for producing substantially lower frictional resistance to motion of the actuator in the direction of such urging than in the opposite direction.

11. The motion-damping assembly of claim having the spring within the housing, the actuator comprising an elongated plunger reciprocable in the housing against the force of the spring, the energy-absorption means including friction surfaces borne by the plunger and the housing and means responsive to the spring to urge the friction surfaces together.

12. The motion-damping assembly of claim 11 having an auxiliary relatively small spring mounted on the plunger and urging the friction surfaces together.

13. In a railway car having a car body rotatably supported on a center bearing of a truck bolster which is spring-supported in a bolster opening of a side frame carried by spaced wheels on a rail the combination of a compressible friction device carried by the car body and slideably engaged with the top of said frame between said wheels when the car body is in normal level position, said friction device comprising means for developing friction during compression at a rate which increases in proportion to the extent of downward movement of the car body from said level position toward said frame, said friction device being releaseable from the frame as the car body rocks upwardly therefrom, and said friction device comprising means for extending itself and for developing friction during gradual extension of said device as the car body rocks upwardly from said frame. 

1. In a railroad car having a wheel-supported side-frame portion, a body portion, and a truck bolster supported by a spring on the side-frame portion and having the body portion pivotally supported thereon, the improved construction having energy-absorbing damping means at each side of the car, bypassing the bolster, for opposing approach motion of that side of the body portion of the car relative to the corresponding side of the side-frame portion of the car but substantially free of resistance to separative motion, the damping means having substantially greater energy-absorption of approach motion in the region of high compression of the spring than in the region of normal compression, and means to disable the damping means during return motion occurring immediately after separative motion substantially faster than critical-speed body-oscillation motion.
 2. The car of claim 1 wherein each damping means includes means primarily responsive to instantaneous relative position of the side-frame portion and the body portion for producing a continuously increasing force characteristic during approach motion.
 3. The car of claim 1 characterized by the damping means having a housing secured to one of said portions and a force-transfer member in sliding engagement with the other of said portions, said disabling means including means to adjust the maximum speed of following of separative motions by the force-transfer member.
 4. The car of claim 1 wherein each damping means comprises a housing, an extending operating member having an inner portion movable in the housing, a spring within the housing, stressed by inward motion of the operating member, adjacent friction surfaces on the housing and the portion of the operating member within the housing, and means within the housing responsive to variation of stressing of the spring to vary the frictional engagement of the friction surfaces correSpondingly.
 5. The car of claim 4 wherein each damping means has an auxiliary adjustable bias spring coupled to the frictional engagement varying means.
 6. The car of claim 5 wherein the bias spring is mounted on the operating member for motion therewith, so that the force exerted thereby is position-independent.
 7. In a railroad car having a track-follower portion and a body portion spring-supported on the track-follower portion for horizontal pivotal motion with respect thereto, the improved construction having energy-absorbing damping means at each side of the car opposing downward motion of that side of the body portion of the car relative to the corresponding side of the track-follower portion of the car but substantially free of resistance to upward motion, the resistance of the damping means to downward motion being substantially greater in the region of high compression of the spring than in the region of normal compression, the damping means having a housing secured to one of said portions and a force-transfer member in sliding engagement with the other of said portions, the force-transfer member comprising a lever pivoted on the same portion to which the housing is secured, the damping means having an extending member reciprocable in the housing and pivotally secured to the lever.
 8. The car of claim 7 wherein the housing is secured to the body portion and the lever has a portion slidingly engaging the side frame of the truck laterally opposite the pivotal support of the body portion.
 9. The car of claim 5 having the housing secured to the end of the body substantially above the bottom thereof, the lever having a portion pivotally mounted on the undersurface of the body and a depending portion slidingly engaging the side frame.
 10. A motion-damping assembly for a railway car having pivotally connected truck and body portions comprising a housing and a fulcrum member both having means for attachment to the same portion of the car, an actuator extending into the housing, a lever pivotally connected to the actuator and the fulcrum member and having a surface constructed and arranged for sliding engagement with an adjacent surface of the other portion of the car, a spring urging the lever into such engagement, and direction-selective energy-absorption means in the housing including means for producing substantially lower frictional resistance to motion of the actuator in the direction of such urging than in the opposite direction.
 11. The motion-damping assembly of claim 10 having the spring within the housing, the actuator comprising an elongated plunger reciprocable in the housing against the force of the spring, the energy-absorption means including friction surfaces borne by the plunger and the housing and means responsive to the spring to urge the friction surfaces together.
 12. The motion-damping assembly of claim 11 having an auxiliary relatively small spring mounted on the plunger and urging the friction surfaces together.
 13. In a railway car having a car body rotatably supported on a center bearing of a truck bolster which is spring-supported in a bolster opening of a side frame carried by spaced wheels on a rail the combination of a compressible friction device carried by the car body and slideably engaged with the top of said frame between said wheels when the car body is in normal level position, said friction device comprising means for developing friction during compression at a rate which increases in proportion to the extent of downward movement of the car body from said level position toward said frame, said friction device being releaseable from the frame as the car body rocks upwardly therefrom, and said friction device comprising means for extending itself and for developing friction during gradual extension of said device as the car body rocks upwardly from said frame. 